Cooling Device

ABSTRACT

Provided is an inexpensive and compact cooling device which does not increase the circulation resistance of a refrigerant, filling quantity of a refrigerant in a natural circulation circuit wherein natural convection of the refrigerant is caused by using a thermo-siphon, and cross-sectional areas of individual passages while maintaining a desired cooling efficiency in the circuit. A secondary cooling device ( 40 ) includes a secondary heat exchange section ( 42 ) of a cascade heat exchanger (HE) which liquifies a gaseous-phase secondary refrigerant, and an evaporator (EP) which vaporizes a liquid-phase secondary refrigerant. The secondary cooling device ( 40 ) is provided with a plurality of natural circulation circuits ( 48 ) equipped with liquid pipings ( 44 ) and gas pipings ( 46 ) which connect the secondary heat exchange section ( 42 ) and the evaporator (EP). The evaporator (EP) has evaporation passages ( 52 ) of the natural circulation circuits ( 48 ) provided in layers vertically apart from one another. The evaporation passage ( 52 ) is formed by a spiral fin tube type heat exchanger whose fins are spirally wound on the outer circumference of an evaporation pipe through which the secondary refrigerant circulates.

TECHNICAL FIELD

The present invention relates to a cooling device equipped with a natural circulation circuit which carries out natural convection of a refrigerant utilizing a temperature gradient between a heat exchange section and an evaporator.

BACKGROUND ART

A cooling device using the thermo-siphon that carries out natural convection of a refrigerant is adopted in a storage facilities, such as a refrigerator, and air conditioners (for example, refer to Patent Document 1). As shown in FIG. 20, a natural circulation circuit 72 includes a condenser 64 which condenses an evaporated refrigerant into a liquefied refrigerant, and an evaporator 66 which is arranged under the condenser 64 to evaporate the liquefied refrigerant to form an evaporated refrigerant, causes the liquefied refrigerant to flow down from the condenser 64 via a liquid piping 68 to the evaporator 66, and circulates the evaporated refrigerant from the evaporator 66 via a gas piping 70 to the condenser 64.

In the condenser 64 and the evaporator 66, a refrigerant which circulates through refrigerant passages 64 a and 66 a provided therein exchanges heat with another medium, such as external air or water, to condense or evaporate the refrigerant. That is, since the cooling efficiency of the cooling device depends on the quantity of heat exchanged between the refrigerant and another medium, the cooling device has the refrigerant passages 64 a and 66 a provided in a meandering manner in the condenser 64 and the evaporator 66 to increase the contact area between the refrigerant passages 64 a and 66 a and another medium (henceforth called heat exchange area).

Used as another form of the evaporator 66 is a so-called fin and tube type heat exchanger that has the refrigerant passage 66 a formed in the meandering manner by using a straight-line piping penetrating a plurality of parallel plate-like fins and a bent portion bent in a U shape to be welded to the end portion of the straight-line piping (for example, refer to Patent Document 1). This fin and tube type heat exchanger is widely adopted as an air-cooling heat exchanger because of its feature to improve the piping density comparatively easily. As another form of an evaporator, there is also known a spiral fin tube type heat exchanger which has a fin spirally wound around the periphery of the evaporation pipe where a refrigerant circulates, and bent in a meandering manner.

Related Art Documents Patent Document

Patent Document 1: Japanese Patent Application Laid-Open No. S63-96463

SUMMARY OF INVENTION Problems to be Solved by the Invention

According to the cooling device, when the piping length required to secure the heat exchange contact area that provides the necessary cooling efficiency is set up, the refrigerant passages 64 a, 66 a become long, increasing the circulation resistance of the refrigerant in the passages 64 a, 66 a. In addition, since the bent portions of the refrigerant passages 64 a, 66 a increase to make the elongated refrigerant passages 64 a, 66 a compact, making the circulation resistance higher. Since the system which uses the thermo-siphon like the cooling device is configured to carry out natural convection of the refrigerant using a temperature gradient between the condenser 64 and the evaporator 66, the circulation power of a refrigerant is weak as compared with the system which carries out enforced circulation of the refrigerant with a pump or the like, so that the smooth flow of the refrigerant is interfered with slight pressure loss or flow resistance against the refrigerant. If the flow of the refrigerant is no longer performed smoothly in the refrigerant passages 64 a, 66 a, the circulation of the refrigerant in the natural circulation circuit 72 including the evaporator 66 becomes worse, or the refrigerant may flow backwards, lowering the delivery capacity of cold, which brings about the problem that the target cannot be cooled efficiently. In order not to reduce the cooling efficiency, therefore, the cooling device needs to set up the cross-sectional area of the refrigerant passages 64 a and 66 a larger according to the circulating amount of the refrigerant, thereby decreasing the circulation resistance of the refrigerant so that the flow state of the refrigerant which is greatly influenced by slight pressure loss is stabilized. However, increasing the diameter of the pipings which constitute the refrigerant passages 64 a, 66 a larger increases restrictions in forming the refrigerant passages, and enlarges the condenser 64 or the evaporator 66, which leads to cost increase.

While the fin and tube type heat exchanger can increase the density of pipings, it has the straight-line piping and the bent portions welded together, so that there are a plurality of welded portions in the refrigerant passages, raising the problem that the reliability against the refrigerant leak is low What is more, since the fin and tube type heat exchanger is configured such that each plate-like fin is connected to every adjoining straight-line piping, if the interval between the individual plate-like fins is narrowed to improve the heat exchanging efficiency, frost is likely to grow at the fins between the adjacent straight-line pipings, and may block the clearance between the pipings, inhibiting the flow of air.

On the other hand, the spiral fin tube type heat exchanger, unlike the fin and tube type heat exchanger, need not perform welding of pipings or expand the pipings during manufacture, significantly simplifying the manufacturing process. In addition, since the spiral fin tube type heat exchanger has no welded portions of the pipings in the circulation passage of the refrigerant, it advantageously has high reliability against refrigerant leak. Further, since the spiral fin tube type heat exchanger is configured to have the fin wound around the evaporation pipe, the fins on the adjacent straight-line pipings are not in contact with each other, so that even if the spiral pitch of the fins is narrowed in order to improve the heat exchanging efficiency, frost, if grown at the fins, does not easily block the clearance between the pipings. That is, the configuration of the spiral fin tube type heat exchanger does not easily cause clogging by frost formation.

However, since the evaporation pipe in the spiral fin tube type heat exchanger is bent after the fin is wound therearound, if the diameter of the evaporation pipe is increased to decrease the circulation resistance of the refrigerant, the minimum bending radius became larger, making the piping density sparse, so that it is difficult to downsize the evaporator. It is therefore difficult to increase the piping density of the evaporation pipe as high as that of the fin and tube type heat exchanger. Accordingly, at present the spiral fin tube type heat exchanger which is excellent in cost, reliability against the refrigerant leak, suppression of frost-originated clogging, etc. over the fin and tube type heat exchanger has not been adopted as the evaporator of the cooling device that uses a thermo-siphon.

The aforementioned configuration for improving the cooling efficiency with a cooling device, and the spiral fin tube type heat exchanger can be adopted by enlarging the heat exchanger. However, one product value of a refrigerator in which a cooling device is used is to increase the internal volume of the refrigerator. To secure this value, it is important to design the heat exchanger (evaporator) in the refrigerator compact, which makes it difficult to enlarge the heat exchanger easily. It is potentially rather demanded to design the heat exchanger (evaporator) in the refrigerator more compact. That is, the cooling device for performing heat exchange efficiently within limited space is actually desired.

Namely, in view of the inherent problems of the cooling devices according to the related art, the present invention has been proposed to overcome the problems properly, and it is an object of the invention to provide an inexpensive and compact cooling device which does not increase the circulation resistance of a refrigerant, filling quantity in a natural circulation circuit wherein natural convection of the refrigerant is caused by using a thermo-siphon, and cross-sectional areas of individual passages while maintaining a desired cooling efficiency in the circuit.

Means for Solving the Problems

To overcome the problems and achieve the desired object, a cooling device according to the invention of the present application, which has a heat exchange section which condenses an evaporated refrigerant, which circulates through a condensation passage, into a liquefied refrigerant, and a tubular evaporation pipe disposed under the heat exchange section to evaporate a liquefied refrigerant which circulates through an internal evaporation passage into an evaporated refrigerant, and is provided with a natural circulation circuit which causes the liquefied refrigerant to flow down to the internal evaporation passage from the condensation passage of the heat exchange section via a liquid piping, and causes the evaporated refrigerant to flow to the condensation passage of the heat exchange section from the evaporation passage via a gas piping, is characterized by comprising an evaporator comprising a set of a plurality of mutually-independent natural circulation circuits and evaporation pipes of the plurality of natural circulation circuits, carbon dioxide being used as a refrigerant which circulates through each natural circulation circuit, and is characterized in that the each evaporation pipe is formed bent in a meandering manner in such a way that a straight portion extends in a transverse direction intersecting a flow direction of air which circulates the evaporation pipes, the liquid piping is connected to a part of the evaporation pipe which is downstream of the flow direction of air, and the gas piping is connected to a part of the evaporation pipe which is upstream of the flow direction of air, and the plurality of evaporation pipes are arranged in layers apart from one another in an up and down relation.

According to the invention, the condensation passage and the evaporation passage are connected by the liquid piping and gas piping so that the natural circulation circuits each constitute a single circuit independently without involving branching of passages and pipings. With the cooling device provided with the natural circulation circuits whose quantity corresponds to the heat exchange contact areas needed in the heat exchange section and evaporator, the condensation passage and the evaporation passage which are needed can be disposed in the heat exchange section and the evaporator, thereby securing the heat exchange contact area needed for the whole circuit. This makes the heat exchange contact area needed per each one of the condensation passage and the evaporation passage smaller, thus making it possible to suppress the required lengths of each condensation passage and each evaporation passage. Shortening the lengths of each condensation passage and each evaporation passage makes the circulation resistance originating from the passage length lower, and reduces the number of meanders, thereby reducing the circulation resistance originating from the bent portions of the passages. As a result, each condensation passage and each evaporation passage can be set up with smaller cross-sectional areas as compared with the related art in which the circulation resistance becomes too large to achieve it, thus making it possible to reduce the amount of the refrigerant which circulates through each condensation passage and each evaporation passage. Since it is possible to reduce the lengths and cross-sectional areas of each condensation passage and each evaporation passage, the heat exchange section and the evaporator can be made compact, reducing the sizes of facilities, such as the capacity of the expansion tank which eases a pressure increase in the circuit. It is therefore possible to make the whole device compact and reduce the cost. In addition, since the individual natural circulation circuits are independent of one another, the drift of the refrigerant is unlikely to occur and it is possible to achieve smooth natural convection of the refrigerant.

Further, with the configuration made to allows the refrigerant to flow upstream from the downstream side of the flow direction of air which circulates through the evaporation pipes, and by using carbon dioxide excellent in heat transfer performance as the refrigerant, it is possible to use an evaporation pipe of a small heat exchange contact area and a small diameter without reducing the heat exchanging efficiency in the evaporator, thus reducing the cost of the evaporator, and making the evaporator more compact. Furthermore, with a plurality of evaporation pipes arranged in layers apart from one another in an up and down relation, these evaporation pipes function as an evaporator with high heat exchanging efficiency.

To overcome the problems and achieve the desired object, a cooling device according to another subject matter of the invention, which has a heat exchange section which condenses an evaporated refrigerant, which circulates through a condensation passage, into a liquefied refrigerant, and a tubular evaporation pipe disposed under the heat exchange section to evaporate a liquefied refrigerant which circulates through an internal evaporation passage into an evaporated refrigerant, and is provided with a natural circulation circuit which causes the liquefied refrigerant to flow down to the internal evaporation passage from the condensation passage of the heat exchange section via a liquid piping, and causes the evaporated refrigerant to flow to the condensation passage of the heat exchange section from the evaporation passage via a gas piping, is characterized in that the natural circulation circuit includes an evaporator comprising a set of a plurality of evaporation pipes and condensation passages equal in number to the a plurality of evaporation pipes, and carbon dioxide is used as a refrigerant which circulates through the natural circulation circuit, a liquid piping which connects to an outflow end of the condensation passage is connected to that evaporation pipe which is different from the evaporation pipe to which a gas piping coupled to an inflow end of the condensation passage is connected, and a gas piping which connects to an outflow end of the evaporation pipe is connected to that condensation passage which is different from the condensation passage to which a liquid piping coupled to an inflow end of the evaporation pipe is concerned, thereby constructing a single natural circulation circuit as a whole, the each evaporation pipe is formed bent in a meandering manner in such a way that a straight portion extends in a transverse direction intersecting a flow direction of air which circulates the evaporation pipes, the liquid piping is connected to a part of the evaporation pipe which is downstream of the flow direction of air, and the gas piping is connected to a part of the evaporation pipe which is upstream of the flow direction of air, and the plurality of evaporation pipes are arranged in layers apart from one another in an up and down relation.

According to the another subject matter, the condensation passage and the evaporation passage are connected by the liquid piping and gas piping so that a single natural circulation circuit as a whole is formed without involving branching of passages and pipings. That is, it is possible to adequately dispose the condensation passages and evaporation passages whose quantities correspond to the heat exchange contact areas needed in the heat exchange section and evaporator. This makes the heat exchange contact area needed per each one of the condensation passage and the evaporation passage smaller, thus making it possible to suppress the required lengths of each condensation passage and each evaporation passage. Shortening the lengths of each condensation passage and each evaporation passage makes the circulation resistance originating from the passage length lower, and reduces the number of meanders, thereby reducing the circulation resistance originating from the bent portions of the passages. As a result, each condensation passage and each evaporation passage can be set up with smaller cross-sectional areas as compared with the related art in which the circulation resistance becomes too large to achieve it, thus making it possible to reduce the amount of the refrigerant which circulates through each condensation passage and each evaporation passage. Since it is possible to reduce the lengths and cross-sectional areas of each condensation passage and each evaporation passage, the heat exchange section and the evaporator can be made compact, reducing the sizes of the facilities, such as the capacity of the expansion tank which eases a pressure increase in the circuit. It is therefore possible to make the whole device compact and reduce the cost. In addition, even with a plurality of condensation passages and a plurality of evaporation passages provided, a single natural circulation circuit as a whole is realized without involving branching of passages and pipings, causing the refrigerant to circulate in good balance through the plurality of condensation passages and the plurality of evaporation passages, so that the drift of the refrigerant does not occur and smooth natural convection of the refrigerant can be achieved.

Further, with the configuration made to allows the refrigerant to flow upstream from the downstream side of the flow direction of air which circulates through the evaporation pipes, and by using carbon dioxide excellent in heat transfer performance as the refrigerant, it is possible to use an evaporation pipe of a small heat exchange contact area and a small diameter without reducing the heat exchanging efficiency in the evaporator, thus reducing the cost of the evaporator, and making the evaporator more compact. Furthermore, with a plurality of evaporation pipes arranged in layers apart from one another in an up and down relation, these evaporation pipes function as an evaporator with high heat exchanging efficiency.

Effect of the Invention

The cooling device according to the invention can be made inexpensive and compact while maintaining the desired cooling efficiency without increasing the circulation resistance of the refrigerant, the amount the refrigerant filled in the circuit, and the cross-sectional area of each passage.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 A side cross-sectional view showing a refrigerator provided with a cooling device according to a preferable first embodiment of the invention as a secondary circuit of a cooling system.

FIG. 2 A schematic circuit diagram showing the essential portions of the cooling system equipped with the cooling device of the first embodiment as the secondary circuit.

FIG. 3 A front view of essential portions showing the evaporation passage according to the first embodiment.

FIG. 4 A top view showing the evaporation passage according to the first embodiment.

FIG. 5 A schematic side view showing the evaporator according to the first embodiment.

FIG. 6 A schematic circuit diagram showing the essential portions of a cooling system provided with a cooling device according to a preferable second embodiment of the invention as a secondary circuit.

FIG. 7 A schematic side view showing an evaporator according to a modification.

FIG. 8 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 9 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 10 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 11 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 12 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 13 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 14 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 15 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 16 A schematic perspective view showing an evaporation pipe according to a modification which is provided with a heat transfer promotion member.

FIG. 17 A schematic perspective view showing an evaporation pipe according to a modification which is not provided with a heat transfer promotion member.

FIG. 18 A schematic perspective view showing an evaporation pipe according to a modification which is not provided with a heat transfer promotion member.

FIG. 19 A schematic side view showing an evaporator according to another modification.

FIG. 20 A schematic diagram showing the natural circulation circuit of the cooling device according to the related art.

DESCRIPTION OF REFERENCE NUMERALS

-   34 primary cooling device (circuit on the primary side) -   42 secondary heat exchange section (heat exchange section) -   44 liquid piping -   46 gas piping -   48 natural circulation circuit -   50 condensation passage -   52 evaporation passage -   52 a inflow end -   52 b outflow end -   55 heat exchanger -   56 evaporation pipe -   56 a straight-line portion -   56C stepped portion -   58 fin (heat transfer promotion member) -   62 natural circulation circuit -   74, 78, 80, 82, 84, 86, 88 heat transfer promotion member -   76 spine fin (heat transfer promotion member) -   EP evaporator

MODES OF CARRYING OUT THE INVENTION

In these days, use of chlorofluorocarbon as a refrigerant is restricted in the system provided with a cooling device, such as a refrigerator or a freezer, from the viewpoint of global warming prevention. In particular, since large-sized systems of business use, such as a freezer machine, use a large amount of chlorofluorocarbon, there are significant demands for reduction of the amount of chlorofluorocarbon used or realization of chlorofluorocarbon-free systems. Accordingly, attention is paid to the secondary loop type refrigeration circuit which has an advantageous circuit configuration in promoting chlorofluorocarbon-free systems. The secondary loop type refrigeration circuit connects two independent circuits, namely a primary-side circuit of a machine compression type which forcibly circulates the refrigerant, and a secondary-side circuit which carries out natural convection of the refrigerant using a thermo-siphon, via a cascade heat exchanger, and can use a heat transfer medium other than chlorofluorocarbon. However, as compared with the refrigeration circuit of the machine compression type which uses chlorofluorocarbon as a refrigerant, the conventional secondary loop type refrigeration circuit has shortcomings of enlarging the whole system, thus requiring a large installation area, and leads to cost increase, and is not competitive to the conventional system using chlorofluorocarbon in terms of the size and price, which stands in the way of promoting the chlorofluorocarbon-free systems. Accordingly, the present inventor has invented the cooling device according to the invention which has a compact and inexpensive configuration, without spoiling the desired cooling efficiency. For example, adapting the cooling device according to the invention to the secondary loop type refrigeration circuit makes it possible to design a system provided with the secondary loop type refrigeration circuit at a size and a cost equivalent to those of the conventional system using chlorofluorocarbon, and overcomes the shortcomings, thereby obtaining the competitive strength on the market. That is, the cooling device according to the invention has an advantageous technical position in promoting the widespread use of the chlorofluorocarbon-free technology based on the secondary loop type refrigeration circuit which is considered important from the viewpoint of global warming prevention. As apparent from the above, the cooling device according to the invention, when adapted to the secondary loop type refrigeration circuit, can overcome the shortcomings of the conventional secondary loop type refrigeration circuit such that it has a large size and is expensive, and can provide generally spreadable technology.

Next, the cooling device according to the invention will be described below by way of a preferred embodiment with reference to the accompanying drawings. The description of the embodiment will be given of a case of employing a so-called secondary loop refrigeration circuit which uses the cooling device according to the invention on the secondary-side circuit, as a large-sized refrigerator, as an example, which is used for business usage as in a shop, and can store a large amount of goods, such as vegetables and meat.

First Embodiment

As shown in FIG. 1, a refrigerator 10 includes a box (adiabatic box) 12 with a thermally insulated structure, which has a storage room 14 internally defined, and a cabinet 16 which is provided above the box 12 and has its outer wall formed by a metal panel 18. An opening 12 a which opens frontward to serve as a goods take-in/take-out port is formed in the box 12 in communication with a storage room 14, and is blocked by a heat insulation door 22 supported at the front portion of the box 12 in an openable/closable manner with a hinge which is not illustrated.

Defined inside the cabinet 16 is a machinery room 20 where a part of a cooling system 32 for cooling the storage room 14 and an electric device box for control (not shown) which controls the cooling system 32 are disposed. Disposed at the bottom of the machinery room 20 is a base plate 24 which is mounted on a top plate 12 b of the box 12 to be a common board for devices disposed in the machinery room 20. An air circulation hole (not shown) that communicates with the machinery room 20 is formed at an adequate location of the metal panel 18 which serves as the outer wall of the cabinet 16, and the atmosphere in the machinery room 20 and the external air interchange each other via the air circulation hole.

A cooling duct 26 is disposed at the upper portion of the storage room 14 away from the top plate 12 b in the box 12 by a predetermined interval, and a cooling room 28 is formed between the cooling duct 26 and the base plate 24 which faces the storage room 14 via a notched port 12 c formed in the top plate 12 b of the box 12. The cooling room 28 communicates with the storage room 14 via a suction port 26 a formed in the front side of the bottom portion of the cooling duct 26, and a cold outlet 26 b formed in the rear side thereof. A blower fan 30 is disposed at the suction port 26 a, so that as the blower fan 30 is driven, air in the storage room 14 is supplied into the cooling room 28 through the suction port 26 a, and cold in the cooling room 28 is supplied to the storage room 14 through the cold outlet 26 b. The notched port 12 c of the top plate 12 b is closed airtight by the base plate 24, so that the storage room 14 (cooling room 28) and the machinery room 20 are partitioned by the base plate 24 to be mutually independent spaces (see FIG. 1).

FIG. 2 is a schematic circuit diagram showing a cooling system 32 equipped with a secondary cooling device (cooling device) 40 according to the first embodiment as a secondary circuit. As shown in FIG. 2, a secondary loop refrigeration circuit in which the primary cooling device (circuit on the primary side) 34 of the machine compression type that forcibly circulates the refrigerant and the secondary cooling device 40 that comprises a thermo-siphon which carries out natural convection of the refrigerant are thermally connected (cascade-connected) in such a way as to exchange heat via a cascade heat exchanger HE is adopted for the cooling system 32. The cascade heat exchanger HE, which is installed in the machinery room 20, includes a primary heat exchange section 36 which constitutes the primary cooling device 34, and a secondary heat exchange section (heat exchange section) 42 which is formed separate from the primary heat exchange section 36 and constitutes the secondary cooling device 40. That is, circuits through which independent refrigerants circulate respectively are formed at the primary cooling device 34 and the secondary cooling device 40. Highly safe carbon dioxide which has characteristics of low viscosity and a high heat transfer coefficient, and which is nontoxic, inflammable and noncorrosive is used as a secondary refrigerant (refrigerant) which circulates through the secondary cooling device 40. In contrast, an HC-based refrigerant, such as butane or propane, or ammonia or the like, which has excellent refrigerant characteristics, such as evaporation heat and saturation pressure, is used as a primary refrigerant which circulates through the primary cooling device 34, and isobutane or propane is used in the first embodiment. That is, the cooling system 32 need not use chlorofluorocarbon as a refrigerant. As the cascade heat exchanger HE, for example, a plate type, a double tube type and a developed type thereof, or equivalents thereto are used.

The primary cooling device 34 is configured by connecting a compressor CM which compresses a gaseous phase primary refrigerant, a condenser CD which liquefies the compressed primary refrigerant, an expansion valve EV which reduces the pressure of the liquid phase primary refrigerant, and the primary heat exchange section 36 of the cascade heat exchanger HE which evaporates the liquid phase primary refrigerant by a refrigerant piping 38 (see FIG. 2). The compressor CM and the condenser CD are disposed commonly on the base plate 24 in the machinery room 20, and a condenser fan FM which carries out forced cooling of the condenser CD is also disposed on the base plate 24 opposite to the condenser CD. In the primary cooling device 34, compression of the primary refrigerant by the compressor CM forcibly circulates the primary refrigerant through the compressor CM, the condenser CD, the expansion valve EV, the primary heat exchange section 36 of the cascade heat exchanger HE, and the compressor CM in the named order, so that the necessary cooling is carried out in the primary heat exchange section 36 under the actions of the individual devices (see FIG. 2).

The secondary cooling device 40 includes the secondary heat exchange section 42 of the cascade heat exchanger HE which liquefies the gaseous phase secondary refrigerant (evaporated refrigerant), and the evaporator EP which comprises a collection of pipings which evaporates the liquid phase secondary refrigerant (liquefied refrigerant), and the secondary heat exchange section 42 and the evaporator EP correspond to each other by 1 to 1 relation (see FIG. 2). The secondary cooling device 40 has a liquid piping 44 and a gas piping 46 which connect the secondary heat exchange section 42 and the evaporator EP, and is provided with natural circulation circuits 48 which supply a liquid phase secondary refrigerant from the secondary heat exchange section 42 to the evaporator EP via the liquid piping 44 under the action of the gravity and circulate a gaseous phase secondary refrigerant from the evaporator EP to the secondary heat exchange section 42 via the gas piping 46. The secondary cooling device 40 of the first embodiment has a plurality of mutually dependent natural circulation circuits 48 (three in the illustrated example, but the quantity can be two or more) formed in parallel. The secondary heat exchange section 42 is disposed in the machinery room 20 while the evaporator EP is disposed in the cooling room 28 (inside the box 12) which is located below the machinery room 20, and is arranged below the secondary heat exchange section 42 with the base plate 24 in between.

The secondary heat exchange section 42 is provided with a plurality of condensation passages 50 (three in the first embodiment) in parallel. The evaporator EP is provided with a plurality of evaporation passage 52 (three in the first embodiment) in parallel. Although the condensation passage 50 is expressed by a straight path from an inflow end 50 a which is connected to the gas piping 46 to an outflow end 50 b which is connected to the liquid piping 44, and the evaporation passage 52 is expressed by a straight path from an inflow end 52 a which is connected to the liquid piping 44 to an outflow end 52 b which is connected to the gas piping 46 in FIG. 2, the condensation passage 50 may be formed meandering or linearly. It is to be noted that the evaporation passage 52 is folded to be meandering as will be described later. In the secondary cooling device 40, a plurality of condensation passages 50, a plurality of evaporation passages 52, a plurality of liquid pipings 44, and a plurality of gas pipings 46 are equal in number. In each natural circulation circuit 48, the liquid piping 44 is laid out to penetrate the base plate 24 with its upper end (start end) connected to the outflow end 50 b of the condensation passage 50 in the secondary heat exchange section 42 while its lower end (termination end) located on the cooling room 28 side is connected to the inflow end 52 a of the evaporation passage 52 in the evaporator EP. In each natural circulation circuit 48, the gas piping 46 is laid out to penetrate with its lower end (start end) located on the cooling room 28 side being connected to the outflow end 52 b of the evaporation passage 52 in the evaporator EP while its upper end (termination end) located on the machinery room 20 side is connected to the inflow end 50 a of the condensation passage 50 in the secondary heat exchange section 42. Reference numeral 54 is a refrigerant charge port provided to fill each natural circulation circuit 48 with the refrigerant.

In the secondary cooling device 40, a temperature gradient is formed in each natural circulation circuit 48 between the secondary heat exchange section 42 which is forcibly cooled by heat exchange with the primary heat exchange section 36 and the evaporator EP, so that the circulation cycle of the refrigerant in which the natural convection of the secondary refrigerant is carried out through the secondary heat exchange section 42, the liquid piping 44, the evaporator EP, and the gas piping 46, then back to the secondary heat exchange section 42 is formed.

As shown in FIG. 3, the evaporation passage 52 comprises a spiral fin tube type heat exchanger 55 which has a fin (heat transfer promotion member) 58 spirally wound around the periphery of an evaporation pipe 56 through which the secondary refrigerant circulates, and is formed by folding the evaporation pipe 56 in a meandering form which has straight portions 56 a and folded portions 56 b (see FIG. 4). The interval of the adjacent straight portions 56 a, 56 a is set in such a way that the fins 58, 58 which are disposed at the individual straight portions 56 a and 56 a are set in such a size as not to contact each other, and the fins 58, 58 are separate from each other. The evaporation pipe 56 is formed so that all the straight portions 56 a and folded portions 56 b are located on one flat surface (hereinafter called “mount surface”), and the mount surface is set to be inclined at a predetermined angle to the horizontal surface (see FIG. 5). The liquid piping 44 is connected to the straight portion 56 a located on the inclined lower end side, the gas piping 46 is connected to the straight portion 56 a located on the inclined upper end side, and the inflow end 52 a of the evaporation passage 52 (evaporation pipe 56) is located below the outflow end 52 b. As shown in FIG. 4, each straight portion 56 a extends in the transverse direction crossing the flow of the air which circulates through the evaporator EP (evaporation pipes 56) and is produced when the blower fan 30 is driven, the inclined upper end side of the mount surface is located upstream of the flow direction of air, and the inclined lower end side is located downstream of the flow direction of air (see FIG. 5). That is, a plurality of straight portions 56 a which constitute the evaporation pipe 56 are arranged so as to shift in order toward the upstream side from the downstream side of the flow direction of the air which circulates through the evaporator ER Therefore, the liquid refrigerant (refrigerant) which has flowed into the evaporation pipe 56 from the liquid piping 44 repeatedly flows into the bottom straight portion 56 a on the downstream side of the flow direction of the air to the straight portions 56 a located upstream sequentially via the folded portions 56 b, and eventually flows out into the gas piping 46 from the topmost straight portion 56 a on the upstream side of the flow direction of air.

As shown in FIG. 5, the individual evaporation pipes 56 are disposed in parallel in layers with the mount surfaces thereof being apart from one another vertically. The fins 58, 58 of the evaporation pipes 56, 56 located one above another are set to be separated from one another. In the first embodiment, the refrigerant inflow end 52 a and the refrigerant outflow end 52 b of each evaporation passage 52 are arranged in such a positional relationship as to be aligned in the vertical direction.

Function Operation of First Embodiment

Next, the function of the cooling system 32 provided with the secondary cooling device 40 according to the first embodiment will be described. In the cooling system 32, when the cooling operation starts, circulation of the refrigerant in each of the primary cooling device 34 and the secondary cooling device 40 will start. First, the primary cooling device 34 will be described. The compressor CM and the condenser cooling fan FM are driven, so that a gaseous phase primary refrigerant is compressed by the compressor CM, primary refrigerant is supplied to the condenser CD via the refrigerant piping 38, and is condensed and liquefied to have a liquid phase by forced cooling by the condenser cooling fan FM. The liquid phase primary refrigerant is depressurized by the expansion valve EV, takes heat (heat absorption) from the secondary refrigerant which circulates through the secondary heat exchange section 42 in the primary heat exchange section 36 of the cascade heat exchanger HE, and is expanded and evaporated at once. The primary cooling device 34 functions in this manner to forcibly cool the secondary heat exchange section 42 by the primary heat exchange section 36 in the cascade heat exchanger HE. Then, the gaseous phase primary refrigerant which has been evaporated in the primary heat exchange section 36 repeats the forced circulation cycle of returning to the compressor CM through the refrigerant piping 38.

In the secondary cooling device 40, since the secondary heat exchange section 42 is cooled by the primary heat exchange section 36, the gaseous phase secondary refrigerant radiates heat to be condensed, changing the phase from the gaseous phase to the liquid phase resulting in an increase in specific gravity, while circulating through each condensation passage 50 of the secondary heat exchange section 42 in each natural circulation circuit 48, the liquid phase secondary refrigerant flows down along each condensation passage 50 of the secondary heat exchange section 42 under the action of gravity. In the secondary cooling device 40, the secondary heat exchange section 42 is disposed in the machinery room 20, and the evaporator EP is disposed in the cooling room 28, located under the machinery room 20, to provide a fall between the secondary heat exchange section 42 and the evaporator ER That is, in each natural circulation circuit 48, the liquid phase secondary refrigerant can be made to naturally flow down, under the action of gravity, towards the evaporator EP which comprises a collection of evaporation pipes 56, via the liquid piping 44 connected to the lower portion of the secondary heat exchange section 42. In the process of circulating through each evaporation passage 52 of the evaporator EP, the liquid phase secondary refrigerant takes heat from the surrounding atmosphere of the evaporator EP, which comprises the collection of evaporation pipes 56, to be vaporized to shift the phase to the gaseous phase. The gaseous phase secondary refrigerant flows from the evaporator EP back to the secondary heat exchange section 42 via the gas piping 46, and in the secondary cooling device 40, the cycle of natural circulation of the secondary refrigerant with easy structure is repeated in each natural circulation circuit 48, without using power, such as a pump or a motor.

The air in the storage room 14 which is sucked into the cooling room 28 from the suction port 26 a by the blower fan 30 is sprayed to the cooled evaporator EP, so that the air which has exchanged heat with the evaporator EP which comprises the collection of evaporation pipes 56 becomes cold. Then, the cold is supplied to the storage room 14 via the cold outlet 26 b from the cooling room 28 to cool the storage room 14. The cold repeats the cycle of circulating through the interior of the storage room 14 and returning into the cooling room 28 via the suction port 26 a. In this case, since the evaporation passage 52 in each natural circulation circuit 48 is configured in such a way that the straight portion 56 a connected to the liquid piping 44 is located on the most downstream side in the flow direction of air, and the straight portion 56 a connected to the gas piping 46 is located on the most upstream side in the flow direction of air, the hot refrigerant exchanges heat with the hot air on the outflow end 52 b side to be vaporized, thus promoting circulation of the secondary refrigerant, so that heat exchange in the evaporator EP is efficiently performed. In addition, the use of carbon dioxide excellent in heat transfer performance as the secondary refrigerant achieves more efficient heat exchange in the evaporator EP which comprises the collection of evaporation pipes 56. That is, the heat exchanging efficiency in the evaporator EP can be improved by making the configuration in such a way that the refrigerant flows upstream from the downstream side of the flow direction of the air which circulates through the evaporator EP, and using carbon dioxide excellent in heat transfer performance as the secondary refrigerant. Therefore, even if a small-diameter pipe with a small heat exchange area is used as the evaporation pipe 56 of the heat exchanger 55, the heat exchanging efficiency in the evaporator EP can be suppressed from falling, thus achieving the compact configuration of the evaporator EP which comprises the collection of evaporation pipes 56. Even if the piping density is reduced by the use of the spiral fin tube type heat exchanger 55, the excellent heat transfer performance of the carbon dioxide used as the secondary refrigerant compensates for the reduction in the heat transfer performance originated from the reduced piping density, thus allowing the spiral fin tube type heat exchanger 55 to appropriately function as the evaporator EP. Further, the refrigerant outflow end 52 b of the evaporation passage 52 is located at a position higher than the inflow end 52 a, it is possible to carry out circulation of the vaporized secondary refrigerant promptly.

The condensation passages 50 and the evaporation passages 52 are connected by the liquid pipings 44 and the gas pipings 46 so that the individual natural circulation circuits 48 each constitute a single circuit independent of one another without involving branching of passages and pipings. Since the individual natural circulation circuits 48 are independent of one another, it is possible to suppress maldistribution of the secondary refrigerant between the condensation passage 50, 50, between the evaporation passage 52, 52, or between the condensation passage 50 and the evaporation passage 52, thus making is possible to match the amount of the secondary refrigerant which circulates through each condensation passage 50 with the amount of the secondary refrigerant which circulates through each evaporation passage 52.

The secondary refrigerant which circulates through each natural circulation circuit 48 may be maldistributed in either the condensation passage 50 or the evaporation passage 52 due to an external factor, such as a change in outside air temperature which acts on the secondary cooling device 40. However, since mutually independent thermo-siphons are formed in the individual natural circulation circuits 48, balance of the secondary refrigerant is automatically adjusted in such a way that the amount of the secondary refrigerant in each condensation passage 50 matches with the amount of the secondary refrigerant in each evaporation passage 52. Therefore, maldistribution of the secondary refrigerant itself is unlikely to occur in each condensation passage 50 and each evaporation passage 52, and even if the maldistribution of the secondary refrigerant occurs, control is performed so that the amount of the secondary refrigerant which circulates through the condensation passage 50 matches with the amount of the secondary refrigerant which circulates through the evaporation passage 52, thus eliminating the need for providing regulation means, such as a valve, to adjust the balance of the secondary refrigerant, and making it possible to simplify the configuration of the secondary cooling device 40. In addition, since smooth natural convection of the secondary refrigerant is carried out in the natural circulation circuit 48, the cooling efficiency in the evaporator EP which comprises the collection of the evaporation pipes 56 can be improved. With the secondary cooling device 40 provided with the natural circulation circuits 48 whose quantity corresponds to the heat exchange areas required for the secondary heat exchange section 42 and the evaporator EP, the condensation passages 50 and the evaporation passages 52 which are needed can be arranged in the secondary heat exchange section 42 and the evaporator EP, thereby securing the heat exchange area needed as the whole system.

In the secondary cooling device 40, a plurality of condensation passages 50 and evaporation passages 52 can be arranged in each of the secondary heat exchange section 42 and the evaporator ER That is, the heat exchange area required of one condensation passage 50 and one evaporation passage 52 becomes smaller, making it possible to shorten the piping length of each condensation passage 50 and each evaporation passage 52. Accordingly, the number of meanders of each condensation passage 50 and each evaporation passage 52, which are made to earn the required piping length can be lessened, and the bent portions which work as the circulation resistance can be reduced, making it possible to reduce the pressure loss of the secondary refrigerant which circulates through the condensation passage 50 and the evaporation passage 52. Since each natural circulation circuit 48 is constituted by a single refrigerant passage without branching the liquid piping 44, the gas piping 46, the condensation passage 50, and the evaporation passage 52, the pressure loss originating from the branched portions of the pipings does not occur. Further, since a head difference of the secondary refrigerant needed for natural convection between the condensation passage 50 and the evaporation passage 52 can be made smaller in each natural circulation circuit 48, a fall needed between the condensation passage 50 and the evaporation passage 52 becomes smaller, it is possible to narrow the vertical interval between the secondary heat exchange section 42 and the evaporator EP, so that the secondary cooling device 40 can be made compact. Since the pressure loss of the secondary refrigerant is small in each natural circulation circuit 48, even if a narrower diameter is selected for the liquid piping 44 and the gas piping 46 as compared with the conventional pipings, the same amount of the secondary refrigerant can be circulated in the circuit, making it possible to reduce the amount of the secondary refrigerant which is filled in the whole circuit.

Since it is possible to reduce the lengths and the cross-sectional areas of each condensation passage 50 and each evaporation passage 52, the secondary heat exchange section 42 and the evaporator EP can be made compact, reducing the sizes of the facilities, such as the capacity of the expansion tank (not shown) which eases a pressure increase in the natural circulation circuit 48, so that the secondary cooling device 40 as a whole can be made compact, achieving cost reduction. In addition, the thickness required to secure the pressure resisting performance of the liquid piping 44, the gas piping 46, and the evaporation pipe 56 can be reduced by decreasing the diameters of these piping 44, 46, and 56. That is, the synergistic effect of narrowing the diameter of each piping 44, 46, 56 and reducing the thickness of each piping 44, 46, 56 can further reduce the weights of the pipings, further reducing the cost. Downsizing of the evaporator EP disposed in the cooling room 28 in the storage room 14 can reduce installation space needed for the evaporator EP (downsizing of the cooling room 28), while the internal volume of the storage room 14 to store goods can be increased, which can improve the commodity value of the refrigerator.

Here, cost reduction by narrowing the diameters of the liquid piping 44, the gas piping 46, and the evaporation pipe 56, etc. will be described specifically. For example, the thickness t of the piping which has pressure resisting performance P is obtained from the following equation where σ is allowable stress of the material and D is the outside diameter of piping.

t=PD/2(σ+P)  (I)

A piping weight M with a length L is obtained by the following equation where C is the specific gravity of the material and D_(i) is the inside diameter of the piping.

M=πLC(D ² −D _(i) ²)/  (II)

Since it is possible to express D_(i)=D−2 t, substituting it to the equation II yields the following equation.

M=πLC(Dt−t ²)  (III)

Substituting the equation I to the equation III yields the following equation.

M=(1−P/2(π+P))×πLCPD ²/2(σ+P)  (IV)

The equation IV shows the weight of the piping which has the pressure resisting performance P. Given that other conditions than D are invariable in the equation IV, the conditions π, L, C, P, and a can be treated as constants. Therefore, the piping weight (the outside diameter D of the piping) which has the pressure resisting performance P can be expressed by the following equation.

M={(1−P/2(σ+P)×πLCP/2(σ+P)}×D ²  (V)

Since inside {} in the equation V is a constant as mentioned above, it can be expressed as M=AD².

The piping weight MD₁ of the piping of the outside diameter D₁ which has the pressure resisting performance P is AD₁ ², and the piping weight MD₂ of the piping of the outside diameter D₂ which has the pressure resisting performance P is AD₂ ². Further, the ratio of the piping weight MD₁ and the piping weight MD₂ is expressed as follows.

MD ₂ /MD ₁ =D ₂ ² /D ₁ ²  (VI)

A description will be given with specific numbers given to the equation VI. In the general cooling device, the outside diameter of the evaporation pipe is often set to 9.52 mm. According to the cooling device of the first embodiment, by way of contrast, the evaporation pipe with the outside diameter of 6.35 mm, which may vary according to the conditions, can be used. Applying these conditions to the equation VI yields the following equation.

MD _(φ)6.35/MD _(φ)9.52=(6.35)²/(9.52)²=0.44

When the evaporation pipe with the outside diameter of 4.76 mm is used in the cooling device of the first embodiment, the following is obtained.

MD _(φ)4.76/MD _(φ)9.52=(4.76)²/(9.52)²=0.25

That is, since the weight ratio of the piping can be said as the ratio of the material price of the piping, it is apparent that according to the secondary cooling device 40 of first embodiment, significant cost reduction can be achieved by narrowing the piping as compared with the conventional cooling device.

Since the diameter of the piping through which the refrigerant circulates can be made narrower by provision of a plurality of natural circulation circuits 48 as mentioned above, enlargement of the evaporator EP can be suppressed even if the spiral fin tube type heat exchanger 55 which has the fin 58 spirally wound around the periphery of the evaporation pipe 56 is used as the evaporation passage 52. That is, reducing the minimum bending radius of the evaporation pipe 56 can make the size compact and can provide a piping density equivalent to that of the fin and tube type heat exchanger. In the spiral fin tube type heat exchanger, it is unnecessary to perform welding of the pipings, or expansion of the pipings in the manufacturing process, so that the manufacturing process is simplified greatly, and the manufacturing cost can be suppressed low. Since the spiral fin tube type heat exchanger 55 has the seamless structure where a U-bent portion is not welded to the linear piping like the fin and tube type heat exchanger, its reliability with respect to leakage of the circulating refrigerant is improved. That is, since the diameters of the pipings in each natural circulation circuit 48 can be reduced, it is possible to use the heat exchanger 55 configured in such a way that the fins 58 at the adjoining linear portions do not contact one another as in the spiral fin tube type with the piping density of the evaporation pipe 56 increased to the level of that of the fin and tube type heat exchanger by the compact configuration with the reduced minimum bending radius of the evaporation pipe 56.

As mentioned above, since the spiral fin tube type heat exchanger 55 has the configuration that does not easily cause clogging originating from frost formation as compared with the fin and tube type heat exchanger, it is possible to lessen the frequency of the defrosting operation to be performed by defrosting means of an unillustrated defrosting heater or the like. This can suppress a temperature rise in the storage room 14 which originates from the frequent defrosting operation. In addition, since the spiral fin tube type heat exchangers 55 which constitute the evaporation passages 52 in a plurality of natural circulation circuits 48 are disposed in parallel layers vertically separated from one another, as shown in FIG. 5, it is possible to make the evaporator EP flat to reduce the ratio of its occupation in the cooling room 28, achieving space reduction. Since the fins 58, 58 of the heat exchangers 55 and 55 arranged in layers vertically are also separated from one another, and clogging originating from frost formation does not easily occur between the heat exchangers 55 and 55.

The cooling system 32 has the primary cooling device 34 and the secondary cooling device 40 connected by the cascade heat exchanger HE, and the primary refrigerant of the primary cooling device 34 and the secondary refrigerant of the secondary cooling device 40 perform heat exchange under evaporation and condensation actions in the cascade heat exchanger HE. That is, since it has a very high heat transfer coefficient as compared with heat exchange only with sensible heat, the heat transfer area between the primary cooling device 34 and the secondary cooling device 40 can be made smaller. Both the primary refrigerant and the secondary refrigerant transfer heat based on latent heat, they can transfer a large amount of heat in a comparatively small amount of them, so that it is possible to reduce the heat capacity of the primary cooling device 34 and the secondary cooling device 40, without reducing the heat exchanging quantity in the cascade heat exchanger HE. Therefore, the amount of the primary refrigerants of the primary cooling device 34 and the amount of the secondary refrigerants of the secondary cooling device 40 can both be reduced, making it possible to reduce the cost and reduce the space for the cooling system 32 due to the compact design of the primary cooling device 34 and the secondary cooling device 40.

Since there is a small amount of the primary refrigerant needed for the primary cooling device 34, the amount of the primary refrigerant can be set equal to or less than the upper limit specified by the law or the like, widening the range of options for the types of refrigerants to be used as the primary refrigerant. Further, the machinery room 20 is set as open space so that air can be changed for the sake of convenience of cooling the condenser CD and the compressor CM with air. Since the primary cooling device 34 is disposed in such a machinery room 20, even if the primary refrigerant should leak out, there is no possibility that it would stay at the machinery room 20. Since the machinery room 20 is partitioned in airtight from the storage room 14 which is closed space by the base plate 24, a leaked primary refrigerant does not flow into the storage room 14, and corrosive gas, such as ammonia or hydrogen sulfide, originating from goods stored in the storage room 14 does not flow into the machinery room 20. In addition, it is possible to choose carbon dioxide excellent in safety as the secondary refrigerant by constituting the cooling system 32 by a secondary loop type refrigeration circuit including the primary cooling device 34 and the secondary cooling device 40. That is, although the evaporator EP faces the storage room 14 (cooling room 28) in the secondary cooling device 40, even if the secondary refrigerant leaks out to the storage room 14, for example, safety of a user can be secured.

The primary cooling device 34 and the secondary cooling device 40 are independent of each other as refrigerant circulating passages, although they are thermally connected by the primary heat exchange section 36 and the secondary heat exchange section 42 of the cascade heat exchanger HE. When the cooling system 32 is stopped (compressor CM: stopped), a hot liquid phase primary refrigerant flows into the primary cooling device 34 from condenser CD at the primary heat exchange section 36. As a result, the temperature of the cascade heat exchanger HE increases, but the temperature of the evaporator EP does not rise since the secondary cooling device 40 is independent, so that a rise in temperature in the storage room 14 at the time of stopping the cooling system 32 is moderate. That is, by cooling the storage room 14 to the necessary preset temperature with the cooling system 32, the time from stopping the cooling system 32 to reactivation of the cooling system 32 can be lengthened. Therefore, the operating ratio of the cooling system 32 is lowered, leading to reduction of the amount of consumed power.

The adaptation of the secondary cooling device 40 of the first embodiment to the cooling system 32 which comprises the secondary loop type refrigeration circuit makes it possible to design the cooling system 32 at a size and cost equivalent to those of the conventional cooling system that uses chlorofluorocarbon, and overcomes the shortcomings of the mechanical compression type refrigeration circuit using chlorofluorocarbon such that it has a large size and need a large installation space, gaining the competitive strength in the market. That is, the secondary cooling device 40 according to the first embodiment holds an effective technical position in promoting the spread of the chlorofluorocarbon-free technology by the secondary loop type refrigeration circuit based on the secondary loop type refrigeration circuit which is considered important from the viewpoint of global warming prevention.

Second Embodiment

FIG. 6 is a schematic circuit diagram showing the cooling system 32 provided with the secondary cooling device (cooling device) 60 according to the second embodiment as the secondary circuit. The cooling system 32 of the second embodiment is installed in the refrigerator 10 described in the description of the first embodiment.

As shown in FIG. 6, the secondary loop refrigeration circuit in which the primary cooling device (circuit on the primary side) 34 of the machine compression type that forcibly circulates the refrigerant and a secondary cooling device (cooling device) 60 that comprises a thermo-siphon which carries out natural convection of the refrigerant are thermally connected (cascade-connected) in such a way as to exchange heat via a cascade heat exchanger HE is adopted for the cooling system 32. Since the configuration of the primary cooling device 34 is the same as that of the first embodiment, the detailed description is omitted and same reference numerals are given to same or like components. With regard to the secondary cooling device 60, same reference numerals are given to those components which are identical or similar to the components of the first embodiment.

The secondary cooling device 60 includes the secondary heat exchange section 42 of the cascade heat exchanger HE which liquefies the gaseous phase secondary refrigerant (evaporated refrigerant), and the evaporator EP which evaporates the liquid phase secondary refrigerant (liquefied refrigerant), and the secondary heat exchange section 42 and the evaporator EP correspond to each other by 1 to 1 relation (see FIG. 6). The secondary cooling device 60 has a liquid piping 44 and a gas piping 46 which connect the secondary heat exchange section 42 and the evaporator EP, and is provided with natural circulation circuits 62 which supply a liquid phase secondary refrigerant from the secondary heat exchange section 42 to the evaporator EP via the liquid piping 44 under the action of the gravity and circulate a gaseous phase secondary refrigerant from the evaporator EP to the secondary heat exchange section 42 via the gas piping 46. The secondary heat exchange section 42 is disposed in the machinery room 20 while the evaporator EP is disposed in the cooling room 28 (inside the box 12) which is located below the machinery room 20, and is arranged below the secondary heat exchange section 42 with the base plate 24 in between. Reference numeral 54 is a refrigerant charge port provided to fill each natural circulation circuit 62 with the refrigerant, and since the secondary cooling device 60 of the second embodiment has a single natural circulation circuit 62, one set of facilities including the refrigerant charge port 54, and a safety valve and an expansion tank (neither illustrated) is sufficient.

The secondary heat exchange section 42 is provided with a plurality of condensation passages 50 (three in the second embodiment; α, β, γ, . . . affixed to the numeral 50 when specifically distinguished) in parallel. The evaporator EP is provided with a plurality of evaporation passage 52 (three in the second embodiment; α, β, γ, . . . affixed to the numeral 52 when specifically distinguished) in parallel. Although the condensation passage 50 is expressed by a straight path from an inflow end 50 a which is connected to the gas piping 46 to an outflow end 50 b which is connected to the liquid piping 44, and the evaporation passage 52 is expressed by a straight path from an inflow end 52 a which is connected to the liquid piping 44 to an outflow end 52 b which is connected to the gas piping 46 in FIG. 6, the condensation passage 47 (sic) may be formed meandering or linearly.

In the secondary cooling device 60, a plurality of condensation passages 50, a plurality of evaporation passages 52, a plurality of liquid pipings 44 (α, β, γ, . . . affixed to the numeral 44 when specifically distinguished), and a plurality of gas pipings 46 (α, β, γ, . . . affixed to the numeral 46 when specifically distinguished) are set equal in number. The liquid piping 44 is laid out to penetrate the base plate 24 with its upper end (start end) connected to the outflow end 50 b of the condensation passage 50 in the secondary heat exchange section 42 while its lower end (termination end) located on the cooling room 28 side is connected to the inflow end 52 a of the evaporation passage 52 in the evaporator EP. The gas piping 46 is laid out to penetrate with its lower end (start end) located on the cooling room 28 side being connected to the outflow end 52 b of the evaporation passage 52 in the evaporator EP while its upper end (termination end) located on the machinery room 20 side is connected to the inflow end 50 a of the condensation passage 50 in the secondary heat exchange section 42.

The secondary cooling device 60 is configured in such a way that the liquid piping 44 connected to the outflow end 50 b of the condensation passage 50 is connected to one of the plurality of evaporation passages 52 that is different from the evaporation passage 52 which is connected with the gas piping 46 coupled to the inflow end 50 a of the condensation passage 50. In the secondary cooling device 60, the gas piping 46 connected to the outflow end 52 b of the evaporation passage 52 is connected to one of the plurality of condensation passages 50 that is different from the condensation passage 50 which is connected with the liquid piping 44 coupled to the inflow end 52 a of the evaporation passage 52, so that one natural circulation circuit 62 as a whole is constituted by a plurality of condensation passages 50, a plurality of evaporation passages 52, a plurality of liquid pipings 44, and a plurality of gas pipings 46. A temperature gradient is formed between the secondary heat exchange section 42 and the evaporator EP which are cooled through heat exchange with the primary heat exchange section 36 which is forcibly cooled in the secondary cooling device 60, forming a circulation cycle in which the refrigerant is naturally convected through the secondary heat exchange section 42, the liquid piping 44, the evaporator EP, and the gas piping 46, and returns to the secondary heat exchange section 42.

The natural circulation circuit 62 formed in the secondary cooling device 60 will be described more specifically with reference to FIG. 6. In the secondary cooling device 60 of the second embodiment, the three condensation passages 50α, 50β, and 50γ are provided in the secondary heat exchange section 42 as the refrigerant passages, and the three evaporation passages 52α, 5β, and 52γ are provided in the evaporator EP as the refrigerant passages. The start end of the first liquid piping 44α is connected to the outflow end 50 b of the first condensation passage 50α, the termination end of the first liquid piping 44α is connected to the inflow end 52 a of the first evaporation passage 52 a, and the secondary liquefied refrigerant is supplied to the first evaporation passage 52α via the first liquid piping 44α from the first condensation passage 50α. The start end of the first gas piping 46α is connected to the outflow end 52 b of the first evaporation passage 52α, the termination end of the first gas piping 46α is connected to the inflow end 50 a of the second condensation passage 50β, and the secondary evaporated refrigerant is returned to the second condensation passage 50β via the first gas piping 46α from the first evaporation passage 52α. The start end of the second liquid piping 44β is connected to the outflow end 50 b of the second: condensation passage 50β, the termination end of the second liquid piping 44β is connected to the inflow end 52 a of the second evaporation passage 52β, and the secondary liquefied refrigerant is supplied to the second evaporation passage 52β via the second liquid piping 44β from the second condensation passage 50β. The start end of the second gas piping 46β is connected to the outflow end 52 b of the second evaporation passage 5213, the termination end of the second gas piping 46β is connected to the inflow end 50 a of the third condensation passage 50γ, and the secondary evaporated refrigerant is returned to the third condensation passage 50γ via the second gas piping 46β from the second evaporation passage 52β. The start end of the third liquid piping 44γ is connected to the outflow end 50 b of the third condensation passage 50γ, the termination end of the third liquid piping 44γ is connected to the inflow end 52 a of the third evaporation passage 52γ, and the secondary liquefied refrigerant is supplied to the third evaporation passage 52γ via the third liquid piping 44γ from the third condensation passage 50γ. The start end of the third gas piping 46γ is connected to the outflow end 52 b of the third evaporation passage 52γ, the termination end of the third gas piping 46γ is connected to the inflow end 50 a of the first condensation passage 50α, and the secondary evaporated refrigerant is returned to the first condensation passage 50α via the third gas piping 46γ from the third evaporation passage 52γ, so that the secondary refrigerant makes one circulation in the natural circulation circuit 62.

Each evaporation passage 52 in the secondary cooling device 60 of the second embodiment comprises the spiral fin tube type heat exchanger 55 shown in FIGS. 3 to 5 as per the first embodiment.

Operation of Second Embodiment

Next, the function of the cooling system 32 provided with the secondary cooling device 60 according to the second embodiment will be described. In the cooling system 32, when the cooling operation starts, the circulation of the refrigerant starts in each of the primary cooling device 34 and the secondary cooling device 60. Since the function of the primary cooling device 34 is described in this specification, paragraph [0029], its description is omitted.

In the secondary cooling device 60, since the secondary heat exchange section 42 is cooled by the primary heat exchange section 36, the gaseous phase secondary refrigerant radiates heat to be condensed, changing the phase from the gaseous phase to the liquid phase, while circulating through each condensation passage 50 of the secondary heat exchange section 42, the liquid phase secondary refrigerant flows down along each condensation passage 50 of the secondary heat exchange section 42 under the action of gravity. In the secondary cooling device 60, the secondary heat exchange section 42 is disposed in the machinery room 20, and the evaporator EP is disposed in the cooling room 28, located under the machinery room 20, to provide a fall between the secondary heat exchange section 42 and the evaporator EP. That is, the liquid phase secondary refrigerant can be made to naturally flow down, under the action of gravity, towards the evaporator EP which comprises the collection of evaporation pipes 56, via the liquid piping 44 connected to the lower portion of the secondary heat exchange section 42. In the process of circulating through each evaporation passage 52 of the evaporator EP, the liquid phase secondary refrigerant takes heat from the surrounding atmosphere of the evaporator EP, which comprises a collection of evaporation pipes 56, to be vaporized to shift the phase to the gaseous phase. The gaseous phase secondary refrigerant flows from the evaporator EP back to the secondary heat exchange section 42 via the gas piping 46, and in the secondary cooling device 60, the cycle of natural circulation of the secondary refrigerant with easy structure is repeated without using power, such as a pump or a motor.

In the natural circulation circuits 62 formed in the secondary cooling device 60, a plurality of condensation passages 50 and a plurality of evaporation passages 52 equal in number to the condensation passages 50 are connected to form a single thermo-siphon which causes the secondary refrigerant to alternately circulate through one condensation passage 50 and one evaporation passage 52. That is, according to the natural circulation circuits 62, a plurality of condensation passages 50 and a plurality of evaporation passages 52 can be provided in a single circuit without branching the liquid piping 44, the gas piping 46, the condensation passage 50 and the evaporation passage 52. Since the natural circulation circuit 62 is thus formed by a single refrigerant passage as a whole, it is possible to suppress maldistribution of the secondary refrigerant between the condensation passages 50, 50, between the evaporation passages 52, 52, or between the condensation passage 50 and the evaporation passage 52, thus making it possible to match the amount of the secondary refrigerant which circulates through each condensation passage 50 with the amount of the secondary refrigerant which circulates through each evaporation passage 52.

The secondary refrigerant which circulates through each natural circulation circuit 62 may be maldistributed in either the condensation passage 50 or the evaporation passage 52 due to an external factor, such as a change in outside air temperature which acts on the secondary cooling device 60. However, since the natural circulation circuit 62 is formed by a single thermo-siphon, balance of the secondary refrigerant is automatically adjusted in such a way that the amount of the secondary refrigerant in each condensation passage 50 matches with the amount of the secondary refrigerant in each evaporation passage 52. Therefore, maldistribution of the secondary refrigerant itself is unlikely to occur in each condensation passage 50 and each evaporation passage 52, and even if the maldistribution of the secondary refrigerant occurs, control is performed so that the amount of the secondary refrigerant which circulates through the condensation passage 50 matches with the amount of the secondary refrigerant which circulates through the evaporation passage 52, thus eliminating the need for providing regulation means, such as a valve, to adjust the balance of the secondary refrigerant, and making it possible to simplify the configuration of the secondary cooling device 60. In addition, since smooth natural convection of the secondary refrigerant is carried out in the natural circulation circuit 62, the cooling efficiency in the evaporator EP can be improved. It is therefore possible to provide a plurality of condensation passages 50 and a plurality of evaporation passages 52 in the secondary heat exchange section 42 and the evaporator EP, so that the heat exchange areas can be obtained without bending or branching the condensation passage 50 and the evaporation passage 52.

In the secondary cooling device 60, a plurality of condensation passages 50 and a plurality of evaporation passages 52 can be arranged in each of the cascade heat exchanger HE and the evaporator ER That is, the heat exchange area required of one condensation passage 50 and one evaporation passage 52 becomes smaller, making it possible to shorten the piping length of each condensation passage 50 and each evaporation passage 52. Accordingly, the number of meanders of each condensation passage 50 and each evaporation passage 52, which are made to earn the required piping length can be lessened, and the bent portions which work as the circulation resistance can be reduced, making it possible to reduce the pressure loss of the secondary refrigerant which circulates through the condensation passage 50 and the evaporation passage 52. Since the secondary cooling device 60 is configured in such a way that the natural circulation circuit 62 is formed by a single refrigerant passage as a whole without branching the liquid piping 44, the gas piping 46, the condensation passage 50, and the evaporation passage 52, the pressure loss originating from the branched portions of the pipings does not occur. Further, since a head difference of the secondary refrigerant needed for natural convection between the condensation passage 50 and the evaporation passage 52 can be made smaller in each natural circulation circuit 62, a fall needed between the condensation passage 50 and the evaporation passage 52 becomes smaller, it is possible to narrow the vertical interval between the secondary heat exchange section 42 and the evaporator EP, so that the secondary cooling device 60 can be made compact. Since the pressure loss of the secondary refrigerant is small in the natural circulation circuit 62, even if a pipe with a narrower diameter is selected for the liquid piping 44 and the gas piping 46 as compared with the conventional pipings, the same amount of the secondary refrigerant can be circulated in the circuit, making it possible to reduce the amount of the secondary refrigerant which is filled in the whole circuit.

Since it is possible to reduce the lengths and the cross-sectional areas of each condensation passage 50 and each evaporation passage 52, the secondary heat exchange section 42 and the evaporator EP can be made compact, reducing the sizes of the facilities, such as the capacity of the expansion tank (not shown) which eases a pressure increase in the natural circulation circuit 62, so that the secondary cooling device 60 as a whole can be made compact, achieving cost reduction. In addition, the thickness required to secure the pressure resisting performance of the liquid piping 44, the gas piping 46, and the evaporation pipe 56 can be reduced by decreasing the diameters of these piping 44, 46, and 56. That is, the synergistic effect of narrowing the diameter of each piping 44, 46, 56 and reducing the thickness of each piping 44, 46, 56, can further reduce the weights of the pipings, and further reduce the cost. Further, even the cooling system 32 according to the second embodiment demonstrates the functional effects described in this specification, paragraphs [0033] and [0038] to [0045].

Since the secondary cooling device 60 of the second embodiment comprises the single natural circulation circuit 62, it is sufficient to provide a single facility such as the refrigerant charge port 54 and a safety valve and an expansion tank (neither illustrated) which prevent an excessive increase in pressure. That is, as compared with the configuration of having a plurality of independent natural circulation circuits 48 as in the secondary cooling device 40 of the first embodiment, the facility becomes compact while maintaining the merits of the first embodiment of preventing drifting of the secondary refrigerant and narrowing the pipe diameter, making it possible to reduce the cost. Since the work of filling the refrigerant in the manufacturing process and maintenance has only to be carried out for the single natural circulation circuit 62 according to the secondary cooling device 60 of the second embodiment, the workability and maintenance work can be improved.

Modifications

The invention is not limited to the configurations of the individual embodiments described above, and may employ other configurations as needed.

1. Although the descriptions of the embodiments have been given of the case where the refrigerant inflow ends and refrigerant outflow ends of a plurality of evaporation passages are arranged in a positional relation to be aligned vertically, the refrigerant inflow ends 52 a and the refrigerant outflow ends 52 b may be arranged to be biased in the flow direction of air, as shown in FIG. 7. 2. Although the descriptions of the embodiments have been given of the case where the evaporation pipe of the spiral fin tube type heat exchanger is arranged in a meandering manner on the same flat surface (mount surface), the straight portions and the folded portions may not be arranged on the same flat surface. For example, as shown in FIG. 19, each evaporation pipe 56 is arranged in a meandering manner on a step-like flat surface, so that the evaporation pipes 56 are formed in a step-like pattern as a whole in the flow direction of the air which circulates through the group of evaporation pipes. More specifically, the evaporation pipe 56 includes an upper step portion 56A which is formed by folding the evaporation pipe 56 in a meandering manner on the flat surface set on the upstream side of the flow direction of the air which circulates through the group of evaporation pipes, a lower step portion 56B which is formed by folding the evaporation pipe 56 in a meandering manner on the flat surface set to a position lower by one step than the upper step portion 56A on the downstream side, and a stepped portion 56C which connects the downstream end of the upper step portion 56A in the flow direction of the air to the upstream end of the lower step portion 56B in the flow direction of the air. The upstream end of the lower step portion 56B extends upstream and from the downstream end of the upper step portion 56A, so that the evaporation pipe 56 is formed in a Z shape as a whole. Those evaporation pipes 56 formed in a step-like pattern are arranged in layers in such a way that the individual steps are apart from one another vertically. Namely, the upper step portion 56A of the evaporation pipe 56 located on the upper side and the upper step portion 56A of the evaporation pipe 56 located on the lower side are arranged in layers, apart from each other vertically, and the lower step portion 56B of the evaporation pipe 56 located on the upper side and the lower step portion 56B of the evaporation pipe 56 located on the lower side are arranged in layers, apart from each other vertically.

The formation of each evaporation pipe 56 in a step-like pattern (Z-shape pattern) with the stepped portion 56C so arranged to have the straight portions 56 a and 56 a overlying each other can make the heat exchange length of each evaporation pipe 56 longer than that in the case where the evaporation pipe is folded in a meandering manner on one flat surface of the same length in the flow direction of air circulating through the group of evaporation pipes. That is, the evaporator EP of the same capability can be constituted by fewer evaporation pipes 56, thus making it possible to reduce the number of parts and the number of assembling steps to reduce the manufacturing cost. In addition, each evaporation pipe 56 is formed in a step-like pattern to rise upward from the inflow end 52 a of the evaporation passage 52 toward the outflow end 52 b, and the straight portions 56 a which form the upper step portion 56A and the lower step portion 56B are formed in such a way as to be biased upward toward the refrigerant outflow end 52 b from the refrigerant inflow end 52 a in order (the flat surface on which the upper step portion 56A and the lower step portion 56B are formed serve as an inclined surface inclined upward toward the outflow end 52 b from the inflow end 52 a), so that the refrigerant which is evaporated in the evaporation pipe 56 is circulated promptly.

Although space where the evaporation pipe 56 does not exist is formed under the upper step portion 56A of the evaporation pipe 56 located on the lower side in the modification shown in FIG. 19, the space portion is where there is little circulation of air, sucked into the cooling room 28 by the blower fan 30, when the evaporator EP is disposed in the cooling room 28 as shown in FIG. 1. That is, even if space is produced by the formation of the evaporation pipe 56 in a step-like pattern, heat exchange with air which flows through the cooling room 28 does not fall significantly. The number of stages of the evaporation pipe 56 is not limited to two, but may be three or more, and the stepped portion 56C which connects the upper step portion 56A and the lower step portion 56B should just be formed so that at least parts of the straight portions 56 a and 56 a have such a relation as to overlie each other vertically.

3. Although the descriptions of the embodiments have been given of the case where the evaporation pipe of the spiral fin tube type heat exchanger is arranged in a meandering manner on the same flat surface (mount surface), the configuration where the straight portion and the folded portion are not arranged on the same flat surface may be employed. For example, the heat exchanger may be configured by arranging the evaporation pipe on the step-like (e.g., crank-shaped) flat surface in a meandering manner, and the individual stages of each heat exchanger are arranged in layers, apart from one another vertically. Even with the configuration where the evaporation pipe is not disposed on the same flat surface, it is preferable to arrange the evaporation pipe in such a way that the linear portion is biased upward toward the refrigerant outflow end side from the refrigerant inflow end side. 4. Although the descriptions of the embodiments have been given of the case where the spiral fin tube type heat exchanger having the fin spirally wound around the periphery of the evaporation pipe is used, this configuration is not restrictive, and configurations shown in FIGS. 8 to 16, for example, can be adopted.

FIG. 8 shows the configuration where the axial (lengthwise direction of the evaporation pipe 56) thickness of a protruding heat transfer promotion member 74 spirally wound around the periphery of the evaporation pipe 56 is increased and the spiral pitch of the heat transfer promotion member 74 is set narrower.

FIG. 9 shows the configuration where the axial thickness of the heat transfer promotion member 74 is set smaller than that of the modification in FIG. 8, and the spiral pitch of the heat transfer promotion member 74 is set wider.

FIG. 10 shows the configuration where multiple spine fins (heat transfer promotion members) 76 are protrusively provided on the periphery of the evaporation pipe 56 spirally (so-called spine fine tube).

FIG. 11 shows the configuration where multiple protruding heat transfer promotion members 78 are protrusively provided on the periphery of the evaporation pipe 56 spirally. The spiral pitch of the heat transfer promotion members 78 according to the modification in FIG. 11, the axial thickness thereof, etc. can be set arbitrarily.

FIG. 12 shows a plurality of plate-like heat transfer promotion members 80 arranged in parallel apart from one another in the lengthwise direction of the evaporation pipe 56. The shape of the heat transfer promotion member 80 shown in the modification in FIG. 12 may be a circle in FIG. 12( a), a quadrangle in FIG. 12( b), an octagon in FIG. 12( c), and a polygon that has projections in the up, down, right and left directions to form a cross as a whole in FIG. 12( d), and may take other various shapes.

FIG. 13 shows plate-like heat transfer promotion members 82 arranged in parallel apart from one another in the lengthwise direction of the evaporation pipe 56 in such a way as to commonly contact a plurality of evaporation pipes 56 (two or more and less than the total number of the evaporation tubes). The shape of the heat transfer promotion member 82 shown in the modifications in FIG. 13 may be a quadrangle in FIG. 13( a), an oval in FIG. 13( b), and a polygon that has projections at the left and right end portions of a rectangular plate in FIG. 13( c), and may take other various shapes.

FIG. 14 shows a plurality of annular heat transfer promotion members 84 arranged in parallel apart from one another in the lengthwise direction of the evaporation pipe 56, with increased axial thickness and wider axial separation intervals. The shape of the heat transfer promotion member 84 shown in the modifications in FIG. 14 may be a circle in FIG. 14( a), a quadrangle in FIG. 14( b), and an octagon in FIG. 14( c), and may take other various shapes.

FIG. 15 shows that a plurality of heat transfer promotion members 86 having a plurality of projections 86 a protrusively provided on the periphery of the evaporation pipe 56 are arranged in parallel apart from one another in the lengthwise direction thereof. The axial separation distance of the heat transfer promotion members 86 according to the modification in FIG. 15, the axial thickness thereof, etc. can be set arbitrarily. With regard to the shape of each projection 86 a, the heat transfer promotion member 86 which comprises a plurality of projections 86 a may have a general shape which is the same as the outer shapes of the heat transfer promotion members 84 of the modifications shown in FIG. 14( b), FIG. 14( c), etc.

FIG. 16 shows the axial thickness of a heat transfer promotion member 88 made larger, and the spiral pitch of the heat transfer promotion member 88 set narrower compared with the modifications in FIG. 14. With regard to the shape of the heat transfer promotion member 88 in the modification in FIG. 16, various shapes can be adopted as in the modifications in FIG. 14. Each heat transfer promotion member 88 may comprise a plurality of projections as in the modification in FIG. 15.

5. Although the descriptions of the embodiments have been given of the case where the spiral fin tube type heat exchanger having the fin spirally wound around the periphery of the evaporation pipe is used, a heat exchange which comprises only an evaporation pipe (pipe body) which is not provided with various kinds of heat transfer promotion members can be adopted since efficient heat exchange in the evaporator is achieved by the arrangement of the evaporation pipe with respect to the flow of air circulating through the evaporation pipe and the use of carbon dioxide excellent in heat transfer performance as the secondary refrigerant. The use of the heat exchange which comprises only an evaporation pipe (pipe body) can make the separation intervals between a plurality of heat exchangers narrower, so that the evaporator can be made more compact.

As the evaporation pipe which is not provided with a heat transfer promotion member, the configurations in FIG. 17 and FIG. 18, for example, can be employed.

FIG. 17 shows the evaporation pipe 56 whose cross-sectional shape is set to a polygonal shape to increase the surface area, and it is possible to take an octagonal shape in FIG. 17( a), a cross shape in FIG. 17( b), and other various shapes.

FIG. 18 shows grooves formed in the top surface of the evaporation pipe 56 to promote heat transfer, and it is possible to take a structure where a series of grooves 90 are formed on the circumference separated in the axial direction as shown in FIGS. 18( a) and 18(b), and a structure where multiple grooves 92 which are not sequential are formed separated in the circumferential direction and the axial direction as shown in FIGS. 18( c) to 18(e). Note that the shapes of the grooves of the modifications shown in FIG. 18 are not limited to the illustrated shapes, and arbitrary shapes can be taken.

6. As the primary cooling device of the cooling system, a refrigeration circuit of an absorption type or refrigeration circuits of other types can also be used. The cooling device according to the invention may be of an air cooling type which cools the heat exchange section with air supplied by a fan or the like. 7. The cascade heat exchanger may have the primary heat exchange section and the secondary heat exchange section formed as separate sections, or may be a heat exchanger of another type. 8. The expansion valve is used in the embodiments as means of depressurizing the liquefied refrigerant in the primary cooling device, which is not restrictive, and a capillary tube or other depressurizing means can be used. 9. The description of the embodiment has been given of the example where the cooling device according to the invention is used on the secondary side of the cooling system which has a secondary loop type refrigeration circuit. Since the cooling device according to the invention can overcome the shortcomings of the cooling system that is provided with a secondary loop type refrigeration circuit as described above, the adaptation of the cooling device according to the invention to the secondary loop type refrigeration circuit is very useful. However, the adaptation of the cooling device according to the invention is not limited to the secondary loop type refrigeration circuit, and the cooling device can be used singularly as a cooling device. 10. The cooling device of the invention can be applied to an air conditioner or the like in addition to so-called storehouses, such as a freezer, a freezing refrigerator, a showcase, and a prefabricated storehouse. 

1. A cooling device comprising a heat exchange section (42) which condenses an evaporated refrigerant circulating through a condensation passage (50) into a liquefied refrigerant, a tubular evaporation pipe (56) disposed under the heat exchange section (42) to evaporate a liquefied refrigerant circulating through an internal evaporation passage (52) into an evaporated refrigerant, and a natural circulation circuit (48) which causes the liquefied refrigerant to flow down to the evaporation passage (52) from the condensation passage (50) of the heat exchange section (42) via a liquid piping (44), and causes the evaporated refrigerant to flow to the condensation passage (50) of the heat exchange section (42) from the evaporation passage (52) via a gas piping (46), wherein the cooling device comprises: a plurality of mutually-independent natural circulation circuits (48) and an evaporator (EP) comprising a set of evaporation pipes (56) of the plurality of natural circulation circuits (48), carbon dioxide being used as a refrigerant which circulates through each natural circulation circuit (48), and wherein the each evaporation pipe (56) is formed bent in a meandering manner in such a way that a straight portion (56 a) extends in a transverse direction intersecting a flow direction of air which circulates the evaporation pipes (56), the liquid piping (44) is connected to a part of the evaporation pipe (56) which is downstream of the flow direction of air, and the gas piping (46) is connected to a part of the evaporation pipe (56) which is upstream of the flow direction of air, and the plurality of evaporation pipes (56) are disposed in layers apart from one another in up and down relation.
 2. A cooling device comprising a heat exchange section (42) which condenses an evaporated refrigerant circulating through a condensation passage (50) into a liquefied refrigerant, and a tubular evaporation pipe (56) disposed under the heat exchange section 02) to evaporate a liquefied refrigerant circulating through an evaporation passage (52) into an evaporated refrigerant, and a natural circulation circuit (62) which causes the liquefied refrigerant to flow down to the evaporation passage (52) from the condensation passage (50) of the heat exchange section (42) via a liquid piping (44), and causes the evaporated refrigerant to flow to the condensation passage (50) of the heat exchange section (42) from the evaporation passage (52) via a gas piping (46), wherein: the natural circulation circuit (62) includes an evaporator (EP) comprising a set of a plurality of evaporation pipes (56) and condensation passages (50) equal in number to the plurality of evaporation pipes (56), and carbon dioxide is used as a refrigerant which circulates through the natural circulation circuit (62), a liquid piping (44) connecting to an outflow end (50 b) of the condensation passage (50) is connected to one of the plurality of evaporation pipes (56) that is different from the evaporation pipe (56) to which a gas piping (46) coupled to an inflow end (50 a) of the condensation passage (50) is connected, and a gas piping (46) which connects to an outflow end (52 b) of the evaporation pipe (56) is connected to one of the plurality of condensation passages (50) that is different from the condensation passage (50) to which a liquid piping (44) coupled to an inflow end (52 a) of the evaporation pipe (56) is connected, thereby constructing a single natural circulation circuit (62) as a whole, the each evaporation pipe (56) is formed bent in a meandering manner in such a way that a straight portion (56 a) extends in a transverse direction intersecting a flow direction of air which circulates the evaporation pipes (56), the liquid piping (44) is connected to a part of the evaporation pipe (56) which is downstream of the flow direction of air, and the gas piping (46) is connected to a part of the evaporation pipe (56) which is upstream of the flow direction of air, and the plurality of evaporation pipes (56) are disposed in layers apart from one another in an up and down relation.
 3. The cooling device according to claim 1, wherein: an inflow end (52 a) to which the liquid piping (44) of the evaporation passage (52) connects is located at a position lower than an outflow end (52 b) to which the gas piping (46) connects.
 4. The cooling device according to claim 1, wherein: the each evaporation pipe (56) is formed in a steplike manner in the flow direction of air circulating the evaporation pipes (56) by a stepped portion (56C) at which the straight portions (56 a) are arranged so as to at least partially overlap each other vertically, and the plurality of evaporation pipes (56) formed in the steplike manner are disposed in layers in such a way that individual steps have an up and down relation.
 5. The cooling device according to claim 1, wherein: heat transfer promotion members (58, 74, 76, 78, 80, 82, 84, 86, 88) are disposed at a periphery of the evaporation pipe (56), and the heat transfer promotion members (58, 74, 76, 78, 80, 82, 84, 86, 88) at adjacent straight portions (56 a) of the evaporation pipe (56) bent in the meandering manner are formed so as to be apart from one another, and the plurality of evaporation pipes (56) are disposed in layers in an up and down relation with the heat transfer promotion members (58, 74, 76, 78, 80, 82, 84, 86, 88) apart from one another.
 6. The cooling device according to claim 1, wherein: the natural circulation circuit (48, 62) is thermally connected via the heat exchange section (42) to a circuit (34) on a primary side of a machine compression type which forcibly circulates a refrigerant.
 7. The cooling device according to claim 2, wherein: an inflow end (52 a) to which the liquid piping (44) of the evaporation passage (52) connects is located at a position lower than an outflow end (52 b) to which the gas piping (46) connects. 